Exemplary embodiments of the present invention relate to a supercritical CO2 generation system for a series recuperative type, and more particularly, to a supercritical CO2 generation system for a parallel recuperative type capable of improving turbine efficiency and saving plumbing costs.
Internationally, as a necessity for efficient power production is increasing more and more and a movement to reduce pollutant emissions is becoming more and more active, various efforts to increase power production while reducing the occurrence of pollutants have been conducted. As one of the efforts, research and development into a power generation system using supercritical CO2 as a working fluid as disclosed in Japanese Patent Laid-Open Publication No. 2012-145092, for example, has been actively conducted.
The supercritical CO2 has a density similar to a liquid state and viscosity similar to gas, such that equipment may be miniaturized and power consumption required to compress and circulate the fluid may be minimized. At the same time, the supercritical CO2 having critical points of 31.4° C. and 72.8 atm is much lower than water having critical points of 373.95° C. and 217.7 atm, and thus may be handled very easily. The supercritical CO2 generation system shows pure generation efficiency of about 45% when being operated at 550° C. and may improve generation efficiency by 20% or more as compared to that of the existing steam cycle and reduce the size of a turbo device.
FIG. 1 is a schematic diagram showing the existing Electronic Power Research Institute (EPRI) proposed cycle.
According to the EPRI proposed cycle of FIG. 1, two turbines 400 are provided and a work of the turbines 400 is transmitted to the compressor 100. The compressor 100 is driven by the work of the turbines to compress a working fluid. The work of the turbines transmitted to the compressor 100 is transmitted to an output corresponding to an output frequency of the generator (not shown) through the gear box (not shown) and transmitted to the generator. A recuperator 200 and heat exchanger 300 using an external heat source, such as waste heat or the like, are provided in plural, and the plurality of recuperators 200 and heat exchangers 300 are arranged in series.
The supercritical CO2 working fluid compressed by the compressor 100 is branched from the first separator S1, and some thereof is transmitted to a low temperature heater 330 and some thereof is transmitted to a low temperature recuperator 230. A working fluid heated by a low temperature heater 330 is transmitted to a first mixer M1. The working fluid transmitted to the low temperature recuperator 230 which exchanges heat with the working fluid transmitted to a pre-cooler 500 is primarily heated and then transmitted to the first mixer M1. The working fluid mixed by the first mixer M1 is transmitted to a second separator S2 where the working fluid is branched and transmitted to a high temperature heater 310 and to a high temperature recuperator 210.
The working fluid transmitted to the high temperature heater 310 is transmitted to a first turbine 410 to drive the first turbine 410 and the working fluid transmitted to the high temperature recuperator 210 that exchanges heat with the working fluid passing through the first turbine 410 is heated and then transmitted to a second turbine 430 to drive the second turbine 430.
The working fluid that is heat-exchanged by the high temperature recuperator 210 through the first turbine 410 and then primarily cooled is transmitted to a second mixer M2, and is mixed with the working fluid passing through a second turbine 430 by the second mixer M2 and transmitted to the low temperature recuperator 230. The working fluid transmitted to the low temperature recuperator 230 exchanges heat with the working fluid branched from the first separator S1 to be secondarily cooled, then transmitted to the pre-cooler 500 to be re-cooled, and then transmitted to the compressor 100.
In the case of the EPRI proposed cycle described above, since the working fluid is introduced by being branched from front ends of the high temperature recuperator 210 and the high temperature heater 310, temperature of transfer pipes 10 and 15 is the same all the times. Therefore, there is a limitation in designing a cycle in which the inlet temperatures of the first turbine 410 and the second turbine 430 are increased to increase a work of the turbine.
Further, there are four transfer pipes 1, 8, 9, and 10 connected to the heat exchanger 300 using an external heat source, which makes it difficult to secure economical efficiency due to the increase in plumbing costs. In addition, since the flow rate of the working fluid mixed by the first mixer M1 is equal to the flow rate of the entire cycle, the first mixer M1 and the pipes before and after the first mixer M1 are relatively large, which leads to increase the plumbing costs.